DE68908181T3 - METHOD FOR OPERATING A COLD STEAM PROCESS UNDER TRANS- OR SUPER-CRITICAL CONDITIONS. - Google Patents

Description

The invention relates to a method for operating a vapor compression circuit; in particular, the invention relates to a method for operating a vapor compression circuit, such as is used in refrigerators, air conditioning units and heat pumps, using a refrigerant operating in a closed circuit under supercritical conditions on the high pressure side.

A conventional vapor compression cycle device for refrigerators, air conditioners or heat pumps is shown in the basic principle in Fig. 1. The device consists of a compressor 1, a condensing heat exchanger 2, a throttle valve 3 and an evaporating heat exchanger 4. These components are connected in a closed circuit in which a refrigerant circulates. The working principle of a vapor compression cycle device is as follows: the pressure and temperature of the refrigerant vapor are increased by means of the compressor 1 before it enters the condenser 2 where it is cooled and condensed, giving up heat to a secondary coolant. The high pressure liquid is then expanded to evaporator pressure and temperature by means of the expansion valve 3. In the evaporator 4 the refrigerant boils and absorbs heat from its surroundings. The vapor at the evaporator outlet is drawn into the compressor, thus completing the cycle.

Conventional vapor compression cycle devices use refrigerants (e.g. R-12, CF2Cl2) that operate entirely at subcritical pressures. A variety of different substances or mixtures of substances can be used as refrigerants. The choice of refrigerant is influenced by, among other things, the condensation temperature, because the critical temperature of the fluid sets the upper limit for the condensation that will occur. To obtain reasonable efficiency, it is usually desirable to use a refrigerant with a critical temperature of at least 20 to 30 K above the condensation temperature. Temperatures close to the critical temperature are usually avoided in the design and operation of conventional systems.

US-A-4205532 describes a heat pump or cooling device in which the refrigerant on the high-pressure side is under supercritical pressure. However, no regulation of the pressure is disclosed.

The relevant technology is covered in detail in the literature, for example in "Handbooks of American Society of Heating, Refrigeration and Air conditioning Engineers Inc.", Fundamentals 1989 and Refrigeration 1986".

The ozone-depleting effect of today's refrigerants (halogenated carbons) has led to significant international efforts to reduce or ban the use of these fluids. Consequently, there is a significant need to develop alternatives to current technology.

In conventional vapor compression cycle devices, capacity control is mainly achieved by regulating the mass flow rate of refrigerant passing through the evaporator. This can be done by controlling compressor capacity, throttling or by bypassing. These methods involve more complicated flow circuits and components, the need for additional equipment and accessories, reduced part load efficiency and other difficulties.

A common type of liquid control device is a thermostatic expansion valve which is controlled by the superheat at the evaporator outlet. Proper operation of the valve under varying operating conditions is achieved by using a significant portion of the evaporator to superheat the refrigerant, resulting in a low heat transfer coefficient.

In addition, the heat release in the condenser of the conventional vapor compression cycle occurs mainly at constant temperature. Therefore, thermodynamic losses due to large temperature differences occur when heat is released to a secondary coolant with a large temperature rise. such as in heat pump applications or when the available flow rate of the secondary coolant is low.

The operation of a vapour compression cycle under transcritical conditions has already been carried out to some extent. Until the time when halogenated carbons became more widespread - about 40 to 50 years ago - CO2 was commonly used as a refrigerant, particularly in ship refrigeration for storage and cargo. The systems were designed to operate normally at subcritical pressures, with evaporation and condensation. Occasionally, typically when a ship was passing through tropical areas, the temperature of the seawater used for cooling could be too high to cause normal condensation and the system was operated under supercritical conditions on the high pressure side. (Critical temperature for CO2 31ºC). In this situation, the amount of refrigerant on the high pressure side was increased to a point where the pressure at the compressor outlet rose to 90 - 100 bar in order to keep the cooling capacity at a reasonable level. The technology of cooling with CO2 is described in older literature, for example by P. Ostertag "Kälteprozesse", Springer 1933 or H.H. MacIntire "Refrigeration Engineering", Wiley 1937.

It was common practice in older CO2 systems to add the necessary supplementary amount from separate storage cylinders. A storage tank installed downstream of the condenser in the normal manner would not be able to provide the functions sought by the present invention.

Another way to determine the capacity and efficiency of a given vapor compression cycle device A method for increasing the pressure of the compressed air, which operates with supercritical pressure on the high-pressure side, is known from the German patent 278095 (1912). This process involves two-stage compression with intermediate cooling in the supercritical range. Compared to the standard system, an additional compressor or pump and an additional heat exchanger must be installed.

The textbook "Principles of Refrigeration" by W.B. Gosney (Cambridge Univ. Press 1982) points out some of the peculiarities of operation near critical pressure. It is suggested that an increase in the amount of refrigerant on the high pressure side may be achieved by temporarily closing the expansion valve, as this will transfer some amount from the evaporator. It is pointed out, however, that this could lead to a shortage of liquid in the evaporator, which in turn would lead to reduced capacity at a time when it is most needed.

It is an object of the present invention to provide a new, improved, simple and effective method for operating a vapor compression cycle under transcritical conditions, by which the specific capacity of the cycle is influenced and which avoids the above deficiencies and disadvantages of the prior art.

Another object of the present invention is to provide a vapor compression circuit which avoids the use of CFC refrigerants and at the same time opens up the possibility of using some refrigerants which are interesting in terms of safety, environmental protection and price.

A further object of the present invention is to provide a new method of capacity control which allows operation at a substantially constant refrigerant flow set and with simple capacity modification by valve operation.

Yet another object of the present invention is to provide a circuit that extracts heat at varying temperatures while reducing heat exchange losses in applications where the secondary coolant flow rate is low or when the secondary coolant must be heated to a relatively high temperature.

The above and other objects are achieved according to the present invention with a method according to claim 1; according to this method, the vapor compression cycle normally operates under transcritical conditions (i.e. supercritical pressure on the high pressure side, subcritical pressure on the low pressure side) where the thermodynamic properties in the supercritical state are used to control the cooling and heating capacity of the device.

According to the present invention, the specific enthalpy at the evaporator inlet is regulated by specifically using the pressure before expansion for capacity control. The capacity is controlled by changing the refrigerant enthalpy difference in the evaporator by changing the specific enthalpy of the refrigerant before expansion. In the supercritical state, the pressure and temperature can be changed independently of each other. According to the invention, this modulation of the specific enthalpy is carried out by changing the pressure before expansion. The refrigerant is cooled as much as possible using the available coolant and the pressure is regulated to achieve the required enthalpy.

The invention is described in more detail below, with reference to the accompanying Figures 1, 2, 3, 4, 5, 6 and 7

Reference is made to the embodiments according to Figures 3 and 4, whereby they are not within the scope of the invention:

Fig. 1 shows a schematic representation of a conventional (subcritical) vapor compression cycle device.

Fig. 2 shows a schematic representation of a transcritical vapor compression cycle device for use in connection with a preferred embodiment of the invention. This embodiment includes a volume as an integral part of the evaporator system which stores refrigerant in a liquid state.

Fig. 3 shows a schematic representation of a transcritical vapor compression cycle device. This embodiment includes an intermediate pressure vessel which is directly integrated into the flow circuit between two valves.

Fig. 4 shows a schematic representation of a transcritical vapor compression cycle device. This embodiment includes a special container that stores refrigerant as a liquid or in the supercritical state.

Fig. 5 is a diagram showing the pressure-enthalpy relationship of the transcritical vapor compression cycle device of Fig. 2, 3 or 4 under different operating conditions.

Fig. 6 shows a set of curves illustrating the control of cooling capacity according to the pressure control method of the present invention. The results shown were measured in a laboratory demonstration system built according to a preferred embodiment of the invention.

Fig.7 shows test results showing the temperature-entropy relationship of the transcritical vapor compression cycle device of Fig.2 operating using carbon dioxide as a refrigerant at different pressures on the high-pressure side.

A transcritical vapor compression circuit for use in the context of the present invention includes a refrigerant whose critical temperature lies between the temperature at which heat is absorbed and the average temperature at which heat is released, and a closed working fluid circuit in which a refrigerant circulates.

Suitable working fluids are, for example: ethylene (Z₂H₄), diborane (B₂H₆), carbon dioxide (CO₂), ethane (C₂H₆) and nitrogen oxide (N₂O).

The closed working fluid circuit consists of a refrigerant flow loop with an integrated storage segment. Fig. 2 shows a preferred embodiment of the invention in which the storage segment is an integral part of the evaporator system. The flow circuit includes a compressor 10 which is connected in series to a heat exchanger 11. a counterflow heat exchanger 12 and a throttle valve 13. The throttle valve can be replaced by an optional expansion device. An evaporating heat exchanger 14, a liquid separator/tank 16 and the low pressure side of the counterflow heat exchanger 12 are connected in flow communication between the throttle valve 13 and the inlet 19 of the compressor 10. The liquid tank 16 is connected to the evaporator outlet 15 and the gas phase outlet of the tank 16 is connected to the counterflow heat exchanger 12.

The counterflow heat exchanger 12 is not essential to the operation of the device, but improves its efficiency, particularly its speed of response to the need for a capacity increase. It also serves to return oil to the compressor. For this purpose, a liquid phase line from the vessel 16 (shown in dashed lines in Figure 2) is connected to the suction line either before the counterflow heat exchanger 12 at 17 or after it at 18 or somewhere between these points. The liquid flow, i.e. refrigerant and oil, is controlled by means of a suitable conventional liquid flow restricting device (not shown in the Figure). By allowing some excess liquid refrigerant to enter the vapor line, a liquid surplus is obtained at the evaporator outlet.

According to Fig. 3, the storage segment of the working fluid circuit includes a container 22 which is integrated into the flow circuit between a valve 21 and the throttle valve 13. The remaining components 10 to 14 of the flow circuit are identical to the components of the previous embodiment, although the heat exchanger 12 can be omitted without major consequences. The pressure in the container 22 is between the pressure on the high pressure side and the pressure on the low pressure side of the flow circuit.

According to Fig. 4, the accumulator segment of the working fluid circuit includes a special accumulator 25 in which the pressure is maintained between the pressure on the high pressure side and the pressure on the low pressure side of the flow circuit. The accumulator segment also consists of valves 23 and 24 which are connected to the high pressure side and the low pressure side of the flow circuit, respectively.

In operation, the refrigerant is compressed in the compressor 10 to a suitable supercritical pressure, with the compressor outlet 20 shown as state "a" in Fig. 5. The refrigerant passes through the heat exchanger 11 where it is cooled to a state "b" giving up its heat to a suitable coolant, for example cooling air or water. If desired, the refrigerant can be further cooled in the counterflow heat exchanger 12 to state "c" before being expanded to state "d". By reducing the pressure in the throttle valve 13, a 2-phase gas/liquid mixture is formed, which is shown by state "d" in Fig. 3. The refrigerant absorbs heat in the evaporator 14 by evaporating the liquid phase.

Starting from the state "e" at the evaporator outlet, the refrigerant vapor can be superheated to the state "f" in the counterflow heat exchanger 12 before entering the compressor inlet 19, thereby completing the loop. In the preferred embodiment of the invention shown in Fig. 2, the state "e" of the evaporator outlet is in the 2-phase region due to the excess liquid at the evaporator outlet.

Modulation of the capacity of the transcritical circuit is achieved by changing the refrigerant state at the evaporator inlet, ie at point "d" in Fig. 5. The cooling capacity per unit of refrigerant flow rate corresponds to the enthalpy difference between states "d" and "e". This enthalpy difference corresponds to a horizontal distance in the enthalpy-pressure diagram, cf. Fig. 5.

The expansion is a constant enthalpy process, which is why the enthalpy at point "d" is equal to the enthalpy at point "c". Consequently, the cooling capacity (in kW) can be controlled at a constant refrigerant flow rate by changing the enthalpy at point "c".

It should be noted that the high pressure single phase refrigerant vapor in the transcritical loop is not condensed in the heat exchanger 11, but is reduced in temperature. The final temperature of the refrigerant in the heat exchanger (point "b") will be several degrees above the temperature of the incoming cooling air or cooling water if counterflow is used. The high pressure vapor can then be cooled several degrees in the counterflow heat exchanger 12 to point "c". The result, however, is that if the inlet temperature of the cooling air or cooling water is constant, the temperature at point "c" will be essentially constant, regardless of the pressure level on the high pressure side.

Thus, the modulation of the device capacity is achieved by changing the pressure on the high pressure side, while the temperature at point "c" remains essentially constant. The course of the isotherms near the critical point causes a change in the enthalpy with pressure, see Fig. 5. The figure shows a reference cycle (abcdef), a cycle with reduced capacity due to reduced pressure on the high pressure side (a'-b'-c'-d'-ef) and a cycle with increased capacity due to higher pressure on the High pressure side (a"-b"-c"-d"-ef). The evaporator pressure is assumed to be constant.

The pressure on the high-pressure side is independent of the temperature because it is filled with a single-phase fluid.

To change the pressure, it is necessary to change the amount of refrigerant on the high-pressure side, i.e. to add or remove parts of the respective refrigerant charge on the high-pressure side. These changes must be compensated by a buffer to avoid liquid overflow or drying out of the evaporator.

In the preferred embodiment for use in connection with the method of the invention, as shown in Fig. 2, the amount of refrigerant on the high pressure side is increased by temporarily reducing the opening of the throttle valve 13. As a result of the correspondingly reduced refrigerant flow to the evaporator, the excess liquid content at the evaporator outlet 15 is reduced. The flow of liquid refrigerant from the tank 16 to the suction line is, however, constant. As a result, the balance between the liquid flow into and out of the tank 16 is shifted, resulting in an overall reduction in the liquid content of the tank and a corresponding accumulation of refrigerant on the high pressure side of the flow circuit.

Increasing the charge on the high pressure side leads to increased pressure and thus to a higher cooling capacity. This mass transport from the low pressure to the high pressure side of the circuit continues until a balance is found between the cooling capacity and the load.

Opening the throttle valve 13 increases the excess liquid content at the evaporator outlet 15 because the amount of evaporated refrigerant is essentially constant. The difference between this liquid flow entering the vessel and the liquid flow from the vessel into the suction line leads to an accumulation. The overall result is a transport of refrigerant charge from the high pressure side to the low pressure side of the flow circuit, reducing the high pressure side charge that is stored in the vessel in liquid state. By reducing the high pressure side charge and thus the pressure, the capacity of the device is reduced until a balance is found.

In addition, a certain amount of fluid transport from the reservoir to the compressor suction line is required to avoid lubricant accumulation in the liquid phase of the reservoir.

According to Fig. 3, the amount of refrigerant on the high pressure side can be increased by simultaneously closing the valve 21 and modulating the throttle valve 13 to provide the evaporator with sufficient liquid flow. This reduces the flow of refrigerant from the high pressure side into the vessel through valve 21, while refrigerant is transported from the low pressure side to the high pressure side by means of the compressor.

A reduction in the filling on the high-pressure side is achieved by opening the valve 21, whereby the flow is kept essentially constant by the throttle valve 13. As a result, mass is transported from the high-pressure side of the flow circuit to the container 22.

According to Fig. 4, the amount of refrigerant on the high pressure side can be increased by opening the valve 24 and simultaneously reducing the flow through the throttle valve 13. As a result, refrigerant charge accumulates on the high pressure side due to the reduced flow through the throttle valve 13. Sufficient refrigerant flow to the evaporator is achieved by opening the valve 24.

A reduction in the high-pressure side charge can be achieved by opening the valve 23 to transport a little refrigerant charge from the high-pressure side to the vessel. Thus, the capacity control is achieved by modulating the valves 23 and 24 and simultaneously operating the throttle valve 13.

The preferred embodiment of Figure 2 has the advantage of simplicity because capacity control is accomplished by operating only one valve. In addition, the transcritical vapor compression cycle constructed according to this embodiment has some ability to self-regulate by adapting to changes in the cooling load by changes in the liquid content of the vessel 16, which results in changes in the high-pressure side charge and hence in the cooling capacity. In addition, operation with excess liquid at the evaporator outlet imparts advantageous heat transfer characteristics.

The device according to Fig. 3 has the advantage of simplified valve operation. The valve 21 only regulates the pressure on the high pressure side of the device, and the valve 13 only ensures the sufficient supply to the evaporator. Therefore, a conventional thermostatic valve can be used for throttling. Oil return to the compressor is achieved simply by allowing the refrigerant to to flow through the container. However, this embodiment does not offer the possibility of capacity control at pressures on the high pressure side below the critical pressure. The volume of the container 22 must be relatively large because it only operates between the discharge pressure and the pressure of the liquid line.

The device of Figure 4 has the advantage of operating as a conventional vapor compression cycle device when running under steady state conditions. The valves 23 and 24 connecting the vessel 25 to the flow circuit are only operated during capacity control. This embodiment requires the use of three different valves during periods of changing capacity.

The latter embodiments have the disadvantage of higher pressure in the vessel compared to the preferred embodiment. However, these differences between the individual systems in terms of design and operating characteristics are not very significant.

Transcritical vapor compression cycle devices built according to the above-described embodiments can find applications in various fields. The technology is ideally suited for small and medium-sized stationary and mobile air conditioning units, small and medium-sized refrigerators/freezers, and small heat pump units. One of the most promising applications is in automotive air conditioning systems, where the current need for a new, CFC-free, lightweight and effective alternative to R12 systems is urgent.

The practical use of the present invention for cooling or heat pump purposes is illustrated by the following examples:

Examples are explained, presenting test results from a transcritical vapor compression cycle device constructed according to the embodiment of Fig. 2 using carbon dioxide (CO2) as the refrigerant.

The laboratory test apparatus uses water as a heat source, i.e. the water is cooled by heat exchange with boiling CO₂ in the evaporator 14. Furthermore, water is used as a coolant, which is heated by CO₂ in the heat exchanger 11. The test apparatus includes a 61 cc reciprocating compressor 10 and a vessel 16 with a total volume of 4 liters. The system further includes a counterflow heat exchanger 12 and a liquid line connection from the vessel to point 17 as shown in Fig. 2. The throttle valve 13 is manually operated.

EXAMPLE 1

This example shows how control of the cooling capacity is achieved by changing the position of the throttle valve 13, thereby changing the pressure on the high pressure side of the flow circuit. By changing the pressure on the high pressure side, the specific refrigerant enthalpy at the evaporator inlet is controlled, resulting in a modulation of the refrigerant capacity at a constant mass flow rate.

The temperature of the water entering the evaporator 14 is kept constant at 20ºC and the temperature of the water entering the heat exchanger 11 is kept constant at 35ºC. The water circulation is constant in both the evaporator 14 and the heat exchanger 11. The compressor runs at a constant speed.

Fig. 6 shows the variation of the cooling capacity (Q), the compressor shaft power (W), the pressure (pH) on the high pressure side, the CO₂ mass flow rate (m), the CO₂ temperature (te) at the evaporator outlet, the CO₂ temperature (tb) at the outlet of the heat exchanger 11 and the liquid level (h) in the vessel when the throttle valve 13 is operated as shown in the top of the figure. The adjustment of the throttle valve position is the only manipulation.

As shown in the figure, the capacity (Q) is controlled in a simple manner by operating the throttle valve 13. Furthermore, it is clear from the figure that under stable conditions the circulating mass flow rate (m) of CO₂ is essentially constant and independent of the cooling capacity. The CO₂ temperature (tb) at the outlet of the heat exchanger 11 is also essentially constant. The curves show that the change in capacity is exclusively a result of the change in pressure (pH) on the high pressure side.

Furthermore, it can be seen from the diagram that the increased pressure on the high pressure side results in a reduction in the liquid level (h) in the vessel, due to the transport of the CO₂ charge to the high pressure side of the circuit.

Finally, it can be observed that the transition period during the capacity increase does not entail significant superheat at the evaporator outlet, i.e. only small fluctuations of te.

EXAMPLE 2

With higher temperature of the water entering the heat exchanger 11 (e.g. higher ambient temperature), the pressure on the high pressure side must be increased in order to maintain a constant cooling capacity. Table 1 shows results of tests carried out at different temperatures (tw) of the water introduced into the heat exchanger 11.

The temperature of the water entering the evaporator is kept constant at 20ºC and the compressor runs at a constant speed.

As the table shows, the cooling capacity can be kept substantially constant as the ambient temperature increases by increasing the pressure on the high pressure side. The refrigerant mass flow rate is substantially constant as shown. Increased pressures on the high pressure side bring about a reduction in the liquid content of the vessel as indicated by the liquid level indications. TABLE 1 EXAMPLE 3

Fig. 8 shows a graphical representation of transcritical cycles in the entropy/temperature diagram. The cycles shown in the diagram are based on measurements on the laboratory test device during operation at five different pressures on the high pressure side. The evaporator pressure was kept constant. The refrigerant was CO2.

The diagram gives a good impression of the principle of capacity control, showing the changes in the specific enthalpy (h) at the evaporator inlet resulting from changes in the pressure (p) on the high pressure side.

Claims (3)

1. Method for operating a vapor compression circuit with a compressor (10), a cooler (11), a throttling device (13) and an evaporator (14) connected in series to form an integral closed circuit operating with supercritical pressure on the high pressure side of the circuit, wherein the pressure on the high pressure side is regulated by changing the respective refrigerant charge on the high pressure side of the circuit by changing the contents of a refrigerant buffer tank arranged in the circuit, wherein the pressure is increased by reducing the contents, and vice versa, thereby influencing the specific capacity of the circuit, and wherein the regulation of the supercritical pressure is carried out by changing the liquid content of a low pressure refrigerant tank (16) located between the evaporator (14) and the compressor (10), using only the throttling device (13) as Control device. 2. A method according to claim 1, characterized in that the condition of the evaporator outlet is maintained as a two-phase mixture of vapor and liquid by creating a liquid excess at the low pressure inlet of an additional heat exchanger (12) where the low pressure refrigerant is subjected to evaporation and superheating by heat from the high pressure refrigerant before inlet to the compressor. 3. Method according to claim 1 or 2, characterized in that the coolant is carbon dioxide.